This invention relates to pressure control valves and particularly to variable load type pressure control valves for railway cars in which a two-stage output gain characteristic is obtained.
In the case of braking railway cars, when brake force F (i.e., F=P.times.f; where P is a pushing pressure of braking shoes, and f is a friction coefficient of brake shoes) exceeds the adhesive force between wheels and rails, the wheels come into a free running condition, commonly referred to as wheel slide. Therefore, the brake force of the cars, F, has been typically set at a value which will not exceed the aforementioned adhesion force, W even in a maximum braking condition, such as an emergency stop.
In general, the friction coefficient of brake shoes f tends to be higher than the adhesion coefficient in low speed operation, even though it is sufficiently lower than the adhesion coefficient in high speed operation. Because of this tendency, the brake force F in maximum braking is normally set at a value which will not exceed the adhesion force W during low speed operation, in order to prevent wheel slide. This results in a problem in which the braking force in high speed operation is substantially less than that which can be supported by the available wheel to rail adhesion. This problem must be addressed, since more powerful brakes are being employed with the development of higher speed cars.
This suggested the possibility of increasing the brake air pressure when train speed increased above a certain speed, in order to increase braking force during high speed operation. One possible solution is to install an extra pressure increase piston into a variable load valve for rail cars which is described in Japanese Patent 62-201557. Such an arrangement is shown in FIG. 3. This variable load valve arrangement of FIG. 3 comprises an air supply chamber 2 connected to a source of compressed air through air supply passage 1; an output chamber 4 connected to an output passage 3; an exhaust chamber 5 opened to atmosphere; a valve seat 7 installed in air supply hole 6 that interconnects the air supply chamber 2 and the output chamber 4; an air supply valve 9 with which a valve spring 8 acts to engage the supply valve with the valve seat 7; an exhaust valve rod 11 inserted in the air supply hole 6 with the front end facing the air supply valve 9 and constituting one open end of a central passage 10, the other end of which opens into the exhaust chamber 5; a control piston 15 including a diaphragm 13 having its outer peripheral edge fixed to the inside of the valve body 12 and its inner periphery fixed to a control piston body 14, a control piston 15 providing the command force that moves the exhaust valve rod 11 in the direction of the air supply valve 9 in response to a command air pressure; a balancing piston 18 which generates a balancing force against the aforementioned command force by receiving the air pressure from the output chamber 4, the balancing piston 18 including a diaphragm 16 having its outer peripheral edge fixed inside the valve body 12 and its inner periphery fixed to the perimeter of the balancing piston 17, an intermediate body 19 having the aforementioned air supply chamber 2, air supply hole 6, valve seat 7, and the air supply valve 9; a position control device 20, which can either move the intermediate body 19 in an axial direction toward the exhaust valve rod 11 or fix it axially in an arbitrary position, a first multiplication member having plural fins 21 that project radially from balancing piston 18 to support diaphragm 16 on one side of the plural fins around the perimeter of the piston body against the aforementioned air pressure effective in output chamber 4, the other side of fin 21 being engageable with control piston 14; a second multiplication member having fins 22 projecting radially from body 12 and arranged within the interstices of the first finned member, the diaphragm supporting surface of the respective fins 21 and 22 being tapered in opposite directions; a first spring 24 supported between a first spring seat 23 that projects through the center of the control piston 15 at one end and by the balancing piston 18 at the other end; a second spring 26 supported by the control piston 15 via a second spring seat 25 at one end and by the balancing piston 18 at the other end; a first spring adjusting mechanism 27 installed between the first spring seat 23 and a threaded portion of the valve body 12 in order to adjust the force of the first spring; a second spring adjusting mechanism 28 installed between the second spring seat 25 and the control piston 15 to adjust the force of the second spring 26.
A pressure increase unit is provided for this variable load valve, as shown in the lower part of the device, comprising a pressure increase piston 31, which applies a pressure increase command force to the aforementioned control piston 15, in response to the supply of command air pressure. Pressure increase piston 31 is formed by a piston body 30 and a diaphragm that is fixed at its outer periphery to valve body 12 and at its inner periphery to piston body 30. In accordance with this arrangement of the pressure increase piston 31, the components of the first spring adjusting mechanism 27 and the second spring adjusting mechanism 28 come to pass through the pressure increase piston body 30. That is to say that the rod for externally operating the positions of the first spring seat 23 and the second spring seat 25 extends through the body 30 downward in the figure. Because of this, O-ring seals 33, 34, and 36 are installed in the places where the rod 32 passes through the control piston body 14, the pressure increase piston body 30, and the valve body 12, respectively.
A variable load valve arranged with the pressure increase piston 31 thus uses the air spring pressure of a railway car as a command air pressure, and operates in the following manner. Since the air spring pressure changes with the car body weight due to loading or unloading of passengers and cargoes, the control piston 15 receives pressure depending on the weight of the car body. The command force caused by the air spring pressure on the control piston 15 is transmitted to the balancing piston 18, and thus to exhaust valve rod 11 to engage and open the air supply valve 9. When the air supply valve opens, the air from the pressure source is supplied to output chamber 4 via air supply chamber 2 and air supply hole 6. When the brake cylinder pressure increases, the pressure of output chamber 4 acting on balancing piston 18 also increases and thus generates the balancing force against the command force of the balancing piston 18. When the balancing force and the command force balance each other, the air supply valve 9 engages valve seat 7 as shown in the figure, while the front end of the exhaust valve rod 11 remains engaged with the air supply valve 9 to achieve a lap condition in which no supply or exhaust of brake cylinder pressure occurs. Therefore, the appropriate output air pressure is obtained in response to a change of the air spring pressure.
The above is a brief explanation of the basic operation of a variable load valve of the type with which the present invention is concerned. It will be noted that, in addition, the effective area ratio of the balancing piston 18 and the control piston 15 can be changed by the structural arrangement related to the intermediate body 19, its position control mechanism 20, the first and second fins 21 and 22, and the diaphragm members 13 and 16. That is to say that the position of the intermediate body 19 can be changed in the upward and downward direction of the figure by rotating the position adjusting mechanism 20. When the position of the intermediate body 19 changes, the axial position at which the aforementioned lap condition occurs also changes, the corresponding position of the balancing piston 18 changes, and finally, the balancing piston effective pressure area changes, since the active area of the balancing piston diaphragm member 16 that is supported by the first fin 21 changes. Because of this, the effective area ratio between the control piston 15 and the balancing piston 18 also changes. The fact that this effective area ratio is changeable, is especially advantageous, since this allows the output air pressure to change in response to a certain command air pressure without having to change the control piston or balancing piston to other different sizes of pistons.
In addition, a desired minimum output air pressure and empty car air spring pressure can be set because of the first spring 24, the second spring 16 and their adjustment control mechanisms 27 and 28, respectively. That is to say that the first spring 24 acts on the balancing piston 18, and this spring force is counteracted by the force of output pressure acting on balancing piston 18. If we suppose that the air spring pressure (command air pressure) is zero, an output air pressure is obtained in response to which the balancing piston 18 acts to balance against the force from the first spring 24. Since the force of the first spring 24 is adjustable by the first spring adjusting mechanism 27, the output air pressure does not become lower than the minimum air pressure as long as the output air pressure is adjusted to the level of the minimum air pressure that is sufficient for the brake cylinder to function when the air spring pressure is zero. That means that the minimum output air pressure can be set by adjusting the first spring 24.
The second spring 26 operates in such a way that it tends to force the balancing piston 18 and the control piston 15 apart. When the spring operating force is small, the second spring is essentially non-existent since the balancing piston 18 and the control piston 15 are engaged with each other, and the command force is transmitted directly to the balancing piston 18, due to the spring compression, even when the air spring pressure (command air pressure) is low. However, if the force of the second spring 26 mentioned above is large, the control piston 15 is disengaged from balancing piston 18, when the air spring pressure is low, and the force of the second spring 26 exerts a counteracting force on the balancing piston 18, causing the situation in which this counteracting force works against the output air pressure on the balancing piston 18. Since the force of the second spring 26 can be adjusted by the second spring adjusting mechanism 28, the output air pressure does not change, due to the absence of position change in the control piston 15, until the air spring pressure exceeds the empty car air spring pressure, by adjusting the force of the second spring 26 in such a way that the control piston 15 is disengaged from balancing piston 18 when the air spring pressure exceeds the empty car air spring pressure. That means that the empty car air spring pressure can be set by adjusting the second spring 26.
By means of the pressure increase piston 31, a pressure increase command is output only during the period the cars run above a certain speed. When pressure is supplied from the car air springs to piston 31, a pressure increase command force is added to the control piston 15 causing it to need a greater balancing force of the balancing piston 18. Therefore, the output air pressure increases. Because of such an arrangement, a larger brake force is achieved in a high speed running condition than in a low speed running condition when brakes are used for running cars.
The conventional variable load valve, explained by FIG. 3 has a fixed pressure amplification ratio depending on the effective pressure area of diaphragm 29 of the pressure increase piston 31. Therefore, there exists a problem in that there needs to be available several pressure increase pistons 31 with different effective pressure areas in cases where another pressure amplification ratio is desired. Furthermore, there is another problem in that the response sensitivity of the output air pressure in response to the command air pressure becomes much worse during high speed running, since the dynamic resistance of the control piston 15 and the pressure increase piston 31 in a pressure increase condition becomes those of not only O-ring seal 33, but also of O-ring seals 34 and 36, whereas, in a condition when there is no pressure increase, as during low speed running, the dynamic resistance is that of only O-ring seal 33. There is also the additional problem that the variable load valve device becomes large because of the large diaphragm size making it necessary to install the pressure increase piston 31, which in turn enlarges the height of the variable load valve device.